1. Field of the Invention
The present invention relates to a scroll compressor comprising an orbiting scroll and a stationary scroll. More particularly, it relates to an improvement in a system for lubricating bearings used for a refrigerant scroll compressor
2. Discussion of Background
The principle of the scroll compressor of this kind will be described with reference to FIG. 11 showing sequential operations of the scroll compressor. FIG. 11 shows the states that a wrap 1a formed on a stationary scroll 1 is combined with a wrap 2a formed on an orbiting scroll 2. The stationary scroll 1 stands still in space, and the orbiting scroll 2 combined with the statioanry scroll 1 has a phase difference of 180.degree. with respect to the stationary scroll 1. The orbiting scroll 2 is caused to revolve around the center O of the stationary scroll, but it does not rotate itself. The states of the stationary and orbiting scrolls at 0.degree., 90.degree., 180.degree. and 270.degree. are respectively shown in FIGS. 11a to 11d. At the angular position of 0.degree. shown in FIG. 11a, a gas-sealing step to confine gas in an intake port 3 is finished, and a compression chamber 5 is formed between the wraps 1a, 2a. As the orbiting scroll 2 revolves, the volume of the compression chamber 5 gradually decreases, and finally, compressed gas is discharged through a discharge port 4 formed at the center of the stationary scroll 1.
FIG. 12 is a longitudinal cross-sectional view of an important portion of a conventional scroll compressor as shown, for instance, in Japanese Patent Application No. 64571/1984. The scroll compressor is used for a hermetic type refrigerant compressor.
In FIG. 12, a reference numeral 1 designates a stationary scroll, immovably placed in space, which has a wrap 1a on a base plate 1b and a dischage port 4 at the center of the base plate 1b, a numeral 2 designates an orbiting scroll which has a wrap 2a on a base plate 2b and a driving shaft 2c projecting from the base plate 2b at the opposite side with respect to the wrap 2a, a numeral 5 designates a compression chamber formed between the wraps 1a, 2a. A main shaft 6 vertically extends in a casing 20. The main shaft fixedly secures the rotor of a motor (not shown) at its lower part, and it has a large diameter part 6a at its top end portion in which an eccentric hole 6b is formed at a position deviated from the axial center of the main shaft, the axial center being designated by an imaginary line 0--0. A bearing 10 is fitted in the eccentric hole 6b and it supports the driving shaft 2c of the orbiting scroll 2 so as to restrict the movement of the driving shaft in the radial direction, whereby an orbiting movement is transmitted to the orbiting scroll 2 when the main shaft 6 is rotated. An oil groove 10a is formed in the bearing 10 along the axial direction. An eccentric oil feeding passage 13 is formed in the main shaft 6 in its axial direction and deviated from the axial center by a distance R.sub.0. A space 18 is formed between the lower end of the driving shaft 2c and the bottom surface of the eccentric hole 6b, the space 18 constituting a second oil pump means 22. In the outer circumferential portion of the large diameter portion 6a of the main shaft 6, an axially extending oil groove 6d is formed. A radial oil feeding conduit 15 is formed passing through the large diameter portion 6a and the bearing 10 to communicate the oil groove 10a formed in the bearing 10 with the vertically extending oil groove 6d, the radial oil feeding conduit 15 constituting a third oil pump means 23. The casing 20 receives therein the stationary and orbiting scrolls 1, 2, the main shaft 6 and the electric motor (not shown). A bearing supporter 7 is fixed to and inside the casing 20. The bearing supporter 7 has a central through hole in which the main shaft 6 is supported through a main bearing 11 so that the main shaft 6 is rotatable, while it is restricted to move in the radial direction. The bearing supporter 7 is also positioned below the orbiting scroll 2 to support the same through a thrust bearing 12. At least one oil groove 12a is formed in the radial direction in a surface of the thrust bearing 12 which supports the orbiting scroll 2 in a slidable manner.
A numeral 8 designates an Oldham coupling which causes an orbital movement of the orbiting scroll 2 while it prevents the rotation of the orbiting scroll 2, a numeral 9 designates a covering plate, and a numeral 14 designates an oil returning passage.
An oil cap 16 with an oil hole 16a at its lower part is attached to the lower end of the main shaft 6, the oil cap constituting a first oil pump means 21 in association with the vertical oil feeding passage 13. A lubricating oil 17 is stored in the bottom of the casing 20 and the oil surface level of the lubricating oil is kept so as to immerse the intermediate portion of the oil cap 16.
A cavity 19 is formed between the upper end portion of the large diameter part 6a and the lower surface of the orbiting scroll 2.
In FIG. 12, a thrust bearing and a radial bearing which are placed below the large diameter part 6a of the main shaft 6 are omitted for simplifying the drawing.
The detail of the thrust bearing 12 is shown in FIG. 14. A plurality of oil grooves 12a are formed in the radial direction in the thrust bearing 12. The width W and the depth H of each of the grooves 12a are made large and the length L is made the shortest by extending in the radial direction so as to reduce the resistance in a passage and increase an amount of oil to be fed. Since the width and the depth of each of the oil grooves 12a are relatively large, it is unnecessary to finish the grooves with high accuracy.
The operation of the scroll compressor having the construction as above-mentioned will be described.
When the main shaft 6 is rotated by actuating the electric motor (not shown), the orbiting scroll 2 starts orbital movement by the aid of the Oldham coupling 8, hence, compression of a refrigerant gas is started. Then, the referigerant gas introduced into the intake port of the casing 20 is sucked into the compression chamber 5 through the inlet port 3 of the stationary scroll 1, compressed in the compression chamber, and finally, is forcibly discharged through the discharge port 4 and a discharge tube (not shown).
The lubricating oil 17 stored in the bottom of the casing 20 is forced into the vertical oil feeding passage 13 by centrifugal pumping action of the first oil pump means 21 and is supplied into a space 18. Then, the lubricating oil is pressurized by the second oil pump means 22; is forcibly passed through the oil groove 10a, and lubricates the bearing 10. The lubricating oil is fed in the radial oil feeding conduit 15 where it is pressurized by the third oil pump means 23 and is retuned to the bottom of the casing 20 through the oil groove 6d in the large diameter part 6a, the cavity 19, the oil grooves 12a, and the oil returning hole 14, during which the lubricating oil lubricates the main bearing 11 and the thrust bearing 10 and the thrust bearing 12 and other parts. The covering plate 9 prevents the lubricating oil 17 from being sucked directly into the intake port 3.
FIG. 9 is a diagram showing the distribution of oil pressure in an oil feeding system constituted by a multi-oil pump means in a conventional compressor. In FIG. 9, the ordinate represents oil pressure and the abscissa represents the position of each part in the oil feeding system. A character A represents the inlet of the oil cap 16, a character B represents the inlet of the vertical oil feeding passage 13, a character C represents the outlet of the vertical oil feeding passage 13, a character D represents the inlet of the oil groove 10a in the space 18, a character E represents the inlet of the radial oil feeding conduit 15, a character F represents the outlet of the radial oil feeding conduit 15 and a character G represents the outlet of the oil groove of the thrust bearing 12. Characters P.sub.1, P.sub.2 and P.sub.3 respectively represent pressure increased by the function of the first and second and third oil pump means 21, 22, 23. Characters .DELTA.P.sub.1, .DELTA.P.sub.2, .DELTA.P.sub.3 respectively represent pressure loss at the outlet of each of the oil pump means.
FIG. 9 shows pressure distribution when a flow rate Q of the oil is .sqroot.2 times as the flow rate Q.sub.1 (which corresponds to the flow rate in an embodiment of the present invention described below). For simplification, the Figure is illustrated on the assumption that the increasing rate of pressure P.sub.1, P.sub.2 or P.sub.3 by each of the oil pump means is constant and the pressure loss .DELTA.P is proportional to the square of the flow rate Q.
FIG. 9 shows occurrence of a negative pressure at points C and E in the oil feeding passage.
In the conventional scroll compressor, the capacity of the pump means at each part in the oil feeding passage is increased and resistance in the oil feeding pass is minimized as much as possible to feed a sufficient amount of oil to each bearing part. However, when the compressor is applied to compress a refrigerant (such as Freon R.sub.12, R.sub.22), the refrigerant is dissolved in the lubricating oil 17. The dissolved refrigerant is gasified when pressure is reduced or temperature is elevated, and decomposition of the lubricating oil 17 takes place. This results in a so-called foaming, whereby the oil feeding pass is closed by the foamed gas to thereby cause great reduction in oil feeding capability.
When the negative pressure is produced in the oil feeding passage, the refrigerant dissolved in the lubricating oil is separated as foamed gas. The foamed gas greatly reduces the oil feeding capability of the second oil pump means 22, and sometimes, it is impossible to supply the lubricating oil. To eliminate the disadvantage described above, the scroll compressor as shown in FIG. 13 has been proposed. In the scroll compressor, a gas vent hole 24 is formed in the main shaft 6 extending downwardly from the bottom of the eccentric hole 6b in its axial direction and having an opening in the outer circumference of the main shaft 6. In this case, the pressure given by the first oil pump means 21 is changed by the gas vent hole 24, and the first oil pump means 21 does not operate in cooperation with the second and third oil pump means 22, 23 under serially connected condition. Namely, when the first oil pump means 21 is stronger, a pattern to weaken the effect of the first oil pump means is generated so that a continuous flow rate is maintained. On the other hand, when the second and third oil pump means 22, 23 are stronger, a pattern of feeding the oil to weaken these pump means is generated around the gas vent hole 24. In more detail, when the second and third oil pump means 22, 23 become stronger, the gas tends to stay at the central portion of the second oil pump means 22 as the space 18 to reduce the centrifugal pumping action of the oil pump means 22. When the gas is further accumulated, the lubricating oil is pushed to the outer diameter portion with respect to the oil groove 10a, and finally, the space 18 is filled with the gas so that only small stream of the lubricating oil is formed from the outlet of the vertical oil feeding passage 13 to the inlet of the oil groove 10a. When the region of the gas expands to the radial oil conduit 15, the action of the pump becomes weak. Thus, when the second and third oil pump means 22, 23 are stronger than the first oil pump means 21, a sufficient lubrication is not obtained because the gas enters into each of the bearing portions.
FIG. 10 shows the distribution of oil pressure in the oil feeding system constituted by the oil pump means in the scroll compressor shown in FIG. 13. The same characters as in FIG. 9 are used in FIG. 10. Since the gas vent hole 19 is formed, a pressure at the outlet (the point C) of the vertical oil feeding passage 13 is substantially equal to a pressure at the intake port (the point A). Accordingly, a flow rate Q.sub.1 given by the first oil means 21 is determined regardless of the characteristics of the second and third oil pump means 22, 23. Similarly, a flow rate Q.sub.2 given by the second and third oil pump means 22, 23 is determined regardless of the first oil pump means 21. The flow rate Q is .sqroot.3/2 times as large as the flow rate Q.sub.2* in FIG. 9. Since pressure loss of the pressurized oil is small in the oil groove 12a of the thrust bearing 12, a negative pressure is produced at the point E (the inlet of the radial oil feeding conduit 15). When Q.sub.2 &gt;Q.sub.1, namely, when the force of the oil pump means 22, 23 is stronger than the first oil pump means 21, there causes reduction in the capacity of the oil pump means 22, 23 so that operation is effected at the point satisfying the condition of Q.sub.2 =Q.sub.1, whereby continuous operating condition is maintained.
When Q.sub.2 &gt;Q.sub.1, a negative pressure is produced at a part of the oil pump means 22, 23. An elevated pressure P.sub.2 by the oil pump means 22 is produced by a centrifugal force acting on the lubricating oil in the space 18 as shown in FIG. 13, and the pressure P.sub.2 is obtainable by the following equation: EQU P.sub.2 =K.sub.2 .gamma./(2g).multidot..omega..sup.2 (R.sub.2.sup.2 -R.sub.1.sup.2) . . . (1)
where .gamma. is the specific weight of the oil, g is acceleration of gravity, .omega. is the angular velocity of the main shaft 6, R.sub.1 is the inner radius of the lubricating oil in a circular form in the space 18, R.sub.2 is the radius of the main shaft 6 extending from the axial center to the inlet portion of the radius oil feeding conduit 15 and K.sub.2 is a coefficient of correction determined in the consideration of continuous distribution of the oil in the cavity.
It is understandable that increase in pressure by the oil pump means 22 becomes small as the radius R.sub.1 becomes large.
Similarly, increase in the pressure P.sub.3 by the oil pump means 23 is expressed by the following equation: EQU P.sub.3 =K.sub.3 .gamma./(2g).multidot..omega..sup.2 (R.sub.3.sup.2 -R.sub.2.sup.2) . . . (2)
where R.sub.3 is the radius of the outer circumference of the large diameter part 6a of the main shaft 6 and K.sub.3 is a coefficient of correction similar to K.sub.2.
As shown in FIG. 10, when the pressure loss .DELTA.P.sub.2, .DELTA.P.sub.3 in the oil pump means 22, 23 are respectively smaller, the inner radius R.sub.1 of the oil in the space 18 becomes large, whereby continuity of the flow rate can be maintained with respect to oil supply from the oil pump means 21. When the radius R.sub.1 becomes large and it is equal to the radius R.sub.2 (R.sub.1 =R.sub.2), the capacity of the oil pump means 22 becomes zero. In this case, when continuity of the flow rate of the oil can not be maintained with respect to the oil supply from the oil pump means 21, the region of gas expands to the radius oil feeding conduit 15, and the value of the radius R.sub.2 becomes substantial in the equation (2). Under the condition, the lubricating oil can not be supplied to each bearing part in a stable manner.
Thus, since the conventional scroll compressor is so constructed that the resistance of the oil feeding passages is made small as possible by increasing the cross-sectional area of the oil groove 12a in the thrust bearing 12, a negative pressure is produced at the inlet portion of the radial oil feeding conduit 15 in the space 18. Accordingly, in the case that the scroll compressor is applied to a refrigerant compressor, the refrigerant in the lubricating oil is gasified to close the oil feeding passage, whereby there may cause damage or seizure of the bearing part.